NASA TECHNICAL NOTE

AERODYNAMIC EVALUATION OF TWO-STAGE AXIAL-FLOW TURBINE DESIGNED FOR BRAYTON-CYCLE SPACE POWER SYSTEM

by Milton G. Kofskey and WilliamJ. Nusbaum

Lewis Research Center CZeveZand, Ohio

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N A T I O N A L A E R O N A U T I C S A N D SPACE A D M I N I S T R A T I O N W A S H I N G T O N , D . C . F E B R U A R Y 1 9 6 8 /

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TECH LIBRARY KAFB, "I

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AERODYNAMIC EVALUATION O F TWO-STAGE AXIAL-FLOW TURBINE

DESIGNED FOR BRAYTON-CYCLE SPACE POWER SYSTEM

By Milton G. Kofskey and William J. Nusbaum

Lewis Resea rch Center Cleveland, Ohio

NATIONAL AERONAUTICS AND SPACE ADMINISTRATION ~~~

For sole by the Cleoringhouse for Federol Scientific ond Technicol lnformotion Springfield, Virginia 22151 - CFSTl price $3.00

AERODYNAMIC EVALUATION OF TWO-STAGE AXIAL-FLOW TURBINE

DESIGNED FOR BRAYTON-CYCLE SPACE POWER SYSTEM

by Milton G. Kofskey and Wi l l iam J. Nusbaum

Lewis Research Center

SUMMARY

An experimental investigation was conducted of a two-stage turbine designed to drive an alternator for a 10-kilowatt-shaft-output sp ace power system. First-stage and two-stage performance resul ts are described for operation at approximately design Reynolds number at equivalent design speed and p res su re ratio with argon as the working fluid. Tes t s were made at speeds ranging from 0 to 120 percent of equivalent design speed and a t p re s su re ratios from 1.02 to 1.65.

The results of the investigation indicated that, for two-stage operation, the static and total efficiencies, based on turbine-inlet and rotor-exit conditions, were 0.825 and 0.845, respectively, which compare favorably with design. The equivalent mass flow, however, was 4 percent lower than design and was attributed to the flow areas in the blade rows being smaller than design. At this point the first- and second-stage total efficiencies were 0.864 and 0.805 or approximately 2 points higher and 4 points lower than design, respectively. A s a result , the stage work split was indicated to be 53 to 47 compared with a design equal split.

Two-stage performance based on turbine -inlet and collector-exit conditions indicated static and total efficiencies of 0.826 and 0.835, respectively. These results, which in­dicate a drop in total efficiency within the collector of approximately 1point with no change in static efficiency, compare closely with design.

INTRODUCTION

NASA is currently investigating components of a Brayton-cycle space power system with a 10-kilowatt-shaft power output. The reference system, described in reference 1, is of the two-shaft arrangement wherein one turbine drives the compressor and another turbine drives the alternator. This arrangement was selected because of the large

difference in speed requirements between the compressor and the alternator and because of problems associated with the use of a speed-reduction gearbox. Rotative speeds in the range of 30 000 to 50 000 rpm a r e required i n order to obtain high compressor effi­ciency with a minimum number of stages. The speed of the alternator is comparatively low at 12 000 rpm because of the system frequency requirement of 400 cycles per second. Therefore, use of the two-shaft arrangement would permit both the compressor and alternator to operate at their optimum rotative speeds without the use of a speed-reduction gearbox.

Cold-performance evaluations have been made on a radial-inflow turbine designed to drive the high-speed compressor used in this reference system. Results of these in­vestigations, as reported in references 2 and 3, indicated that efficiencies in the 85- to 90-percent range can be obtained for the compressor-drive turbine.

A two-stage axial-flow turbine with a mean diameter of 8 . 5 inches (21.59 cm) was designed to drive the alternator. The turbine was designed and fabricated under contract, and a description of the design and mechanical testing of the turbine by the contractor is given in reference 4.

Results of Reynolds number investigations using axial-flow turbines show that the turbine efficiency level is dependent on Reynolds number. For example, the resu l t s of reference 5 show a 5-point decrease in turbine efficiency when the Reynolds number was decreased f rom 1.2X106 to 1.08X10 5 . Since the subject alternator turbine was designed to operate at a Reynolds number of 4. 95x104 , which is substantially lower than that normally encountered, it was considered important to establish the performance charac ­t e r i s t ics of this turbine at these low-Reynolds-number conditions.

Accordingly, an experimental investigation of the two-stage turbine was made at Lewis to determine the overall and stage performance of this unit. The turbine package used conventional oil-lubricated bearings to simplify the performance tes ts , instead of the gas bearings which would be used fo r the space power system.

Turbine performance t e s t s were made with argon as the working fluid at an inlet temperature of 610° R (339' K) and at an inlet p re s su re of 2 .5 psia (1.723 newtons/cm 2) to correspond to design Reynolds number at design speed and p res su re ratio. Data were obtained over a range of total- to s ta t ic-pressure ratios of 1.02 to 1.65 and over a range of speeds from 0 to 120 percent of design value.

This report presents design information and turbine performance for first-stage and two-stage operation. Test resu l t s are presented in t e r m s of equivalent specific work, torque, mass flow, and efficiency. A radial survey of rotor-exit flow angle and total p ressure is also presented for both first-stage and two-stage operation.

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TUR B INE DESCR IPTION

A s mentioned in the INTRODUCTION, the two-stage axial-flow turbine was designed to drive a low-speed alternator for a 10-kilowatt-shaft-output space power system with argon as the working fluid. A detailed description of the aerodynamic and mechanical design of this turbine is given in reference 4. For convenience, the more important design values and turbine features are presented herein. The design point values are

TABLE I. - TURBINE DESIGN VALUES

Operation

First-stage Two-stage ~

1685 (936.11) 1685 (936.11) 8.45 (5.826) 8.45 (5.826) 0.611 (0.277) 0.611 (0.277)

12 000 12 000

_____- - - - -_ 1.2540 1.1206 1.2542

- - - - -_-____ 1.2495

Iesign point (argon) Inlet total temperature, T;, OR (OK)

h l e t total pressure, p i , psia (newtons/cm 2) M a s s flow, w, lb/sec (kg/sec) Turbine rotative speed, N, rpm Total- to static-pressure ratio

Overall, pi/p6 Rotor exit, pi/p3 or

Total- to total-pressure ratio Overall, pi/pg Rotor exit, p;/p; or

Blade-jet speed ratio, v Total to static efficiency, q s

Overall Rotor exit

Total to total efficiency, qt Overall Rotor exit

1.1144 1.2469 0.651 0.465

-----_----_ 0.826 0.805 0.825

_ _ _ _ _ _ _ _ _ _ _ 0.843 0.847 0.850

15.04 (35.010) 49 500

1. 537 (0.697)

5.96 (13.874) 54.06 (6.108) 108.12 (12.216)

7554 7554

- - -_ -__-___ 1.226

1.106 1.225

- -_______-_ 1.232 1.112 1.233 0.651 0.465

3

Specific work, Ah, Btu/lb (joules/gram) Reynolds number, Re = w / p r m

iir equivalent

M a s s flow, EW ,Tl6, lb/sec (kg/sec)Ocr

Specific work, Ah/@,,, Btu/lb (joules/gram)

7.52 (17.505) 49 500

1. 537 (0.697)

2.98 (6.937) Torque, T E / ~ ,in. -lb (newton-meter) Rotative speed, N/ficY, rpm Total- to total-pressure ratio

Overall, Pi/& eq

Rotor exit, Pi/P; or 5,eq Total- to static-pressure ratio

Overall, p'1/p 6 ,eq Rotor exit, P ; /P~ or 5,eq

Blade-jet speed ratio, v

12.0

lVIV I = 0.094"1

10.4

42. I

Stdtion

(u/v,,i = n. 239 IUfV,,I 3

* 0.280

1111 Mt?dn. IO 110

Figure 1. - Design velot.ity didgrdnis.

listed in table I. Design-point values are given at design inlet p re s su re and temperature with argon as the working fluid. A i r equivalent values for the design point are also given with the assumption of the same values of efficiency as for argon.

Velocity Diagra rns

From the design conditions of pressure, temperature, weight flow, and speed, veloc­ity diagrams were calculated to meet the design work requirement, with the assumption of an equal work split between the two stages. Values of efficiency which vere used in the calculation included the effects of both low Reynolds number and rotor blade tip leakage. The predicted efficiencies were based on loss coefficients which were adjusted in accordance with the 1/5 power law fo r viscous losses. Velocity diagrams (fig. 1) were calculated at the hub, mean, and tip diameters to give free vortex flow patterns. The diagrams indicate a conservative aerodynamic design with a relatively low level of velocity through both stages. The turning at the mean diameter is 67.3' and 67.8' for the first- and second-stage rotors , respectively, with a large amount of reaction in both stages. Stator-exit angles of approximately 72' at the mean diameter are also quite conservative, in that exit angles of 75' o r less are usually considered favorable for low blade leaving losses. There is a small amount of turbine-exit whirl in the direction of rotation. Prerotation of the flow at the turbine inlet, as shown in figure 1, simulates that condition which would be established by an axial-flow compressor drive turbine, described in reference 6, which is also par t of this technology program.

First-stage stator

First-stage rotor

CD-9057 (a) Hub. (b) Mean. (cl Tip.

Figure 2. - Stator and rotor blade passages and profiles.

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Stator Blading

The stator design resulted in 44 blades for the first stage and 40 blades for the second stage with respective values of solidity at the mean diameter of 1.57 and 1.56. The stator blade passages and profiles at the hub, mean, and tip diameters are shown in figure 2. The stators are quite conventional with a convergence of the flow passage from the inlet to the throat that results, generally, i n accelerating flow in that portion of the passage on both the suction and pressure surfaces. The design blade-surface velocity distributions at the hub, mean, and tip diameters are presented in reference 4. The figures therein show only a small amount of deceleration on the suction surface downstream of the throat and on the pressure surface near the leading edge.

Rotor Blading

The rotor design incorporated 36 blades in each of the two stages with tip diameters of 9.7 and 9.8 inches (24.6 and 24.9 cm) for the first and second stages, respectively. The resulting values of solidity at the mean diameter were 1.3 and 1.4. A low solidity (1.2) at the tip diameters together with a large amount of blade twist, as dictated by the velocity diagrams, resulted in a short guided channel in the tip region of the blade. This distinctive feature is seen in figure 2, which shows the rotor blade passages and profiles

-3770

Figure 3. - Two-stage t u r b i n e rotor assembly.

6

ter

23.5 in. (59.7 cm)

CD-9272 (a) Two-stage operation.

Figure 4. - Cross section of turb ine.

7

I

n.

Exhaust collector--, . 23.5 in. ~

(59.7 cm)

(b) First-stage operation.

=I CD-9273

Figure 4. - Concluded.

fo r both rotors. The large amount of blade twist is apparent in figure 3, which shows the rotor assembly. Rotor blade taper in both the axial chord and blade thickness, as shown in figure 2, is conducive to low blade stresses. The design blade-surface velocity distributions at the hub and mean diameters are presented in reference 4. There is gen­erally accelerating flow from the inlet to the throat of both rotors with only a small amount of deceleration on the suction surface downstream of the throat. The decel­erations were limited to small values in the blading design to increase the assurance of obtaining good blade performance, particularly in the low-Reynolds-number region encountered in this application.

T u r b i n e Assembly

Figures 4(a) and (b) show cross-sectionalviews of the turbine for the two-stage and first-stage operations, respectively. The major dimensions of the turbine a r e given, as well as the arrangement of the blading, the gas flow passages, the bearings, and other details of the mechanical design. Included are 36 inlet guide vanes designed to give a prerotation to the flow of about 12'at the mean diameter, as stipulated by the velocity diagrams. Three s t ru ts afforded support for the inlet duct inner wall and also served as a path for the tubing connected to the static pressure taps in the inner wall.

Rotor blade tip clearances were 0.012 and 0.015 inch(0.030 and 0.038 cm) for the f i r s t and second stages, respectively. These values were approximately 1.0 and 1.1 percent of the blade heights. Both figures (4(a) and (b)) show the exhaust collector which includes the sharp radially outward turn in the flow passage immediately downstream of the blading. The rotating assembly w a s supported by a roller bearing and a ball bearing, which are jet-oil lubricated.

APPARATUS, INSTRUMENTATION, AND PROCEDURE

The apparatus consisted of the turbine, described in the preceding section, an air-brake dynamometer to absorb and measure the power output of the turbine, and an inlet and exhaust piping system with flow controls. The arrangement of the apparatusis shown schematically in figure 5. Pressurized argon was used as the driving fluid for the turbine. The argon was piped into the turbine through an electric heater, a fi l ter , a weight-flow measuring station consisting of a calibrated flat-plate orifice, and a remotely controlled pressure-regulating valve. The gas, after passing through the turbine, was exhausted through a system of piping and a remotely operated valve into the laboratory low-pressure exhaust system. With a fixed inlet pressure, the remotely operated value in the exhaust line was used to obtain the desired pressure ratio ac ross the turbine.

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Argon from h igh­(7­pressure supply system

heater Pressure Flat-plate cont ro l valveor i f ice

To low-pressure exhaust 1 system and control valves 1

supply system

CD-9274

Figure 5. - Experimental equipment.

The airbrake dynamometer, which was cradle mounted on air bearings for torque measurement, absorbed and measured the power output of the turbine and a t the same time, controlled the speed. The force on the torque a r m was measured with a c o m ­mercial strain-gage load cell. The rotational speed was measured with an electronic counter in conjunction with a magnetic pickup and a shaft-mounted gear. The turbine test facility is shown in figure 6.

The instrument measuring stations are shown in figure 4. Overall performance (flange to flange) f o r the two-stage operation (fig. 4(a)) was based on measurements taken at stations 1 and 6. Turbine performance was also determined by measurements taken at stations 1 and 5 (exit of the second-stage rotor). The following instrumentation was located at the turbine inlet (station 1): eight static p re s su re taps (four each on the inner and outer walls), a self-alining probe for flow angle measurement, and two total-temperature rakes (each containing three thermocouples). At station 5, immediately downstream of the second-stage-rotor trailing edge, the instrumentation consisted of eight static pressure taps (four each at the inner and outer walls) and a self-alining probe for flow angle, total-temperature, and total-pressure measurements. In the calcu­lation of pressure ratio a c r o s s the turbine, pressures at stations 1and 5 were determined from the average of the static pressures a t the inner and outer walls. Four static pres­s u r e taps and two total-temperature rakes were located at station 6. The temperature rakes were used primarily as a check on turbine efficiency as calculated from torque measurements. Interstage pressure taps were installed to determine the variation of

10

Figure 6. - Experimental t u rb ine test setup.

static pressure through the turbine. Four static pressure taps were located in the outer wall at the exit of the first-stage stator (station 2), first-stage rotor (station 3), and second-stage stator (station 4).

Instrumentation for the first-stage operation (fig. 4(b)) was similar to that for the two-stage operation, except that stations 4 and 5 were eliminated and additional instru­mentation was installed at station 3. In this case, performance was determined from measurements taken at station 1 and at the exit of the first-stage rotor (station 3) . Instrumentation a t station 3 for the first-stage operation consisted of eight static p re s ­su re taps (four each at the inner and outer walls) and a self-alining probe for flow angle, total-temperature, and total-pressure measurements. P r e s s u r e ratio ac ross the stage was determined from the average of the pressures measured at the inner and outer walls.

The absolute values of the p re s su res at the various stations were measured by the use of manometer tubes which contained a fluid of low specific gravity (1.04) and were evacuated to 10.0 microns of mercury (0.00013 newtons/cm 2) on tne reference side of the manometer tube. A l l other data were recorded by an automatic digital potentiometer and processed through an electronic digital computer.

For both the two-stage and single-stage modes of operation, performance data were taken at nominal inlet totalconditions of 610.2' R (339' K) and 2. 5 psia (1.723 newtons/ cm2). These values of temperature and pressure correspond to a Reynolds number of about 47 400 at equivalent design speed and pressure ratio. This value of Reynolds

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number is near the design value of 49 500. Reynolds number, as used herein, is defined as Re = w/prm . For the two-stage operation, data were taken over a range of equiva­lent inlet-total- to exit-static-pressure ratio (Pi/p5) from about 1.08 to 1.66 and over a range of speeds from 0 to 120percent of equivalent design speed. For the first-stage operation, data were taken over the same range of speed but at the slightly different range of equivalent p re s su re ratios of 1.05 t o 1.56. For both modes of operation, a rotor-exit radial survey was made of flow angle, total pressure, and total temperature at equivalent design speed and pressure ratio.

Friction torque of the bearings and seals was obtained by measuring the amount of torque required to rotate the shaft and rotor over the range of speeds covered in the investigation. In measuring the friction torque, windage losses were minimized by evacuating the air from the turbine. A friction torque value of approximately 1.0 inch-pound (0.113 newton-meters) was obtained at equivalent design-point operation. This value corresponds to about 5.5 percent of two-stage turbine torque obtained at equivalent design rotative speed. Friction torque was added to shaft torque when turbine efficiency was determined.

The aerodynamic performance of the two-stage turbine was determined on the basis of both total and static efficiency from measurements taken at the turbine inlet (station 1) and second-stage rotor exit (station 5) (fig. 4(a)). P r e s s u r e s a t these stations were obtained from an average of the static p re s su res at the inner and outer walls. Overall performance of the turbine with exhaust-collector was obtained and then compared with that of the turbine in order to determine the effectiveness of the exhaust collector. The overall performance was based on both total and static efficiencies and was obtained from measurements taken at stations 1and 6. First-stage performance was obtained from measurements taken as stations 1and 3 (fig. 4@)) where pressures at both stations were determined from an average of the static p re s su res at the inner and outer walls. The total pressures were calculated from weight flow, static p re s su re , total temperature, and flow angle from the following equation:

In the calculation of total pressure at station 6, the flow was assumed to be normal to the plane defined by that station.

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I

RESULTS AND DISCUSSION

The results of this investigation are presented in two sections covering first-stage and two-stage performance, respectively. This performance w a s investigated at approximately design Reynolds number (at equivalent design speed and pressure ratio) with argon as the working fluid. A l l data, with the exception of the radial surveys, are shown i n t e rms of equivalent air values. Experimental design-point performance results are given in table II, along with the values assumed in the design, to aid in the following discussion of results for both first-stage and two-stage performance.

TABLE II. - PERFORMANCE VALUES

Operation I First-stage Two-stage I

~_______

Design Experimental De sign ~

Reynolds number, Re 49 500 46 900 49 500 47 700

Total efficiency,;vt, 1 to 3 (or 5) 0.847 0.867 850 0.845 Static efficiency, os, to (or 5) 0.805 0.827 0.825 0.825 Overall total efficiency, vt, to - - - _ - _ _ _ _ _ 0.843_ _ _ _ - _ _ _ _ _ 0.835 Overall static efficiency, q s , to - _ _ _ _ _ _ - _ _ 0.826_ _ _ _ - _ _ _ _ _ 0.826 Equivalent specific work, 2.98 (6.94) 3.07 (7. 15) 5.96 (13.87) 5.96 (13.87)

Ah/8,,, Btu/lb (joules/gram) Equivalent mass flow, 1.537 (0.697) 1.468 (0.666) 1.537 (0. 697) 1. 477 (0. 670)

E W B c r 6 , lb/sec (kg/sec)</Equivalent torque, T E / ~ ,in. -1b 54.06 (6. 11) 53.99 (6. 10) 108.12 (12.22) 104.44 (11.80)

L (newton-meters)

aBased on rotor-exit conditions. bBased on total efficiencies of 0.847 for first and second stages.

First-Stage Per formance

Figure 7 shows the variation of the first-stage turbine specific work Ah/Bcr with total- to static-pressure ratio p1/p3 for lines of constant speed. The specific work obtained at design speed and p res su re ratio was 3.07 Btu p e r pound (7.15 joules/gram). This value is 3 percent greater than the design value of 2.98 Btu pe r pound (6.94 joules/ gram). The low design pressure ratio of 1.112 indicates that the turbine had low subsonic internal velocity levels. The figure also shows that limiting loading w a s not reached over the range of pressure ratios investigated.

8In figure 8, mass flow EW G/6 is plotted against exit-static- to inlet-total­

13

I .

u

VI

c W

1.0 1.1 1.2 1.3 1.4 1.5 1.6 Equivalent inlet-total - to exit-static­

pressure ratio, p i /p3

Figure 7. -Va r ia t i on of specific work wi th pressure rat io and speed for first-stage operation.

1. I z. 4

1.0- 2.2.

U

-: . 9 - z - 2.0 m Y L o ­

1

@ ,g-@1.8

EE 3-1.6­0

-g- . 7 - = VI Percent of int L VI VI equivalent -E 1.4- design speed2 c- . 6 - h

m 0 120-m .% 1.2- 0 100 ­.->

CT3 CT . 5 - - 0 90 W D 70 -

1.0- a 50 )

. 4 - F 30

I \ . 3 L . 6” I

.60 .65 .70 1. Equivalent exit-static- to inlet-total-pressure ratio, p3/p\

Figure 8. -Var ia t ion of mass flow wi th pressure rat io and speed for first-stage operation.

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pressure ratio p3/pi for lines of constant speed. At the design pressure ratio of 0. 90, the m a s s flow was 1.468 pounds pe r second (0.666 kg/sec). This value is 4 percent lower than the design value of 1.537 pounds per second (0.697 kg/sec). Measurements of the first-stage stator and rotor throat areas indicated them to be smaller than the design values by 8 and 4 percent, respectively. Therefore, the deficiency in m a s s flow may be attributed to the smaller throat areas. The trends of the m a s s flow curves (i. e. , the variation of mass flow with pressure ratio for lines of constant speed) are typical of subsonic turbines. The curves show that, for any given p res su re ratio, mass flow increases with increasing speed.

The variation of torque 7 ~ / 6with speed for lines of constant pressure ratio p;/p3 is shown in figure 9. These resul ts are obtained from faired data because the

Percent of equivalent design speed

Figure 9. -Va r ia t i on of torque wi th speed and pressure ra t io for first-stage operation.

data were taken at constant blade speeds and not at constant p re s su re ratios. The torque at design speed and pressure ratio (1.112) was 53.93 inch-pounds (6.10 newton-meters) which is essentially equal to the design value of 54.06 inch-pounds (6.108 newton-meters). The attainment of design torque, coupled with the 4-percent decrease in mass flow, resulted in the increase in turbine specific work. The figure shows that zero speed torque, at design p res su re ratio, is 112.40 inch-pounds (12.7 newton-meters), which is 2.1 t imes the value obtained at design speed and pressure ratio.

Figures 10 (a) and (b)present the static and total efficiencies of the first stage over the range of blade-jet speed rat ios investigated. These figures show that the maximum efficiency occurs in the region of the design blade-jet speed ratio of 0.651. A static

15

0 - 0

7 I

(a) Static efficiency.

. 3 .5 . 7 Blade-jet speed ratio, u (b) Total efficiency.

iPerceni

eq uiva Ient design speed­

120 100 ­

90

50 30 -

I I

1. 1

Figure 10. -Var ia t ion of eff iciency w i th blade-jet speed rat io for first-stage operation.

16

x

efficiency of 0.827 and a total efficiency of 0.867 were obtained at design speed and design total- to static-pressure ratio. These values are approximately 2 points higher than the respective design values of 0.805 and 0.847.

A radial survey of first-stage rotor-exit total p re s su re and rotor-exit flow angle was taken at design speed and p res su re ratio. Figure 11 (a) shows the variation of exit total p re s su re with radius ratio. The dashed line represents design total- to total-pressure ratio. A region of comparatively low total pressure occurs near the hub at a radius ratio of about 0.77. Small deviations of total pressure from design value were obtained over the greater portion of the passage height. Hub and tip static p res su res are plotted in the figure to indicate the low velocity level.

Figure 11 (b) shows large deviations in exit flow angle from design. The variation in flow angle and the low total p re s su re in the region near the inner wall suggest that this may be a core of low-momentum fluid. This fluid comes from the first stator in the region of the blade suction surface and the inner wall o r from a redistribution of low-momentum accumulations in the rotor. Since the stator and rotor throat areas were smaller than design and the measured throat dimension did not vary in a uniform manner

.91

3 0 1 1 ma= 201 : 0­

1a--m 101 1-.­w I

01 i .72 .76

--D

(a) E x i t total pressure.

.80 ,84 .88 .92 .96 1.00 Radius ratio, r/router wall

(b) Exit flow angle.

F igure 11. -Var ia t i on of rotor-exit total pressure and flow angle wi th radius rat io for first-stage operation at equivalent design speed and design pressure ratio.

17

from hub to tip, some deviations of exit flow angle over the passage height would also be expected.

Figure 11 (b) also shows that there was underturning (as indicated by positive exit angles) over most of the radial distance. The combination of underturning at the rotor exit and the attainment of specific work which was la rger than design indicates that the whirl component at the rotor inlet was la rger than design. This condition results from the stator blade angles being more tangential than design and in turn accounts for the flow area being smaller than design.

Two-Stage Per fo rmance

A s mentioned in APPARATUS, INSTRUMENTATION, AND PROCEDURE, two-stage turbine performance was determined by measurements taken at stations 1 and 5 (turbine inlet and rotor exit) and at stations 1 and 6 (turbine inlet and exit of collector). Effi­ciency based on measurements at stations 1 and 5 gives the performance of the turbine to the rotor exit. Efficiency based on measurements at stations 1and 6 gives an overall turbine performance which is of importance for the space power system since the exit collector is par t of the turbine package for the system. Blade-jet speed ratio, in all

O L I I I I 1 0 I . 1 1. 2 1 .3 1.4 1.5 1 .6 1.7

Equivalent inlet-total- to exit-stat ic-pressure ratio, Pi lpg

Figure 12. -Var ia t i on of specific work w i th pressure rat io and speed for two-stage operation.

18

cases, was based on conditions at the turbine inlet and the rotor exit. The variation of specific work Ah/OCr with pressure ratio p;/p5 and speed is

shown in figure 12. A specific work value of 5.96 Btu per pound (13.87 joules/gram) was obtained at design speed and pressure ratio. This value agrees with the design value of 5.96 Btu pe r pound (13.87 joules/gram). For the speed range of 30 to 120 percent of design, turbine work was constant in the region of design pressure ratio, which indicates a constant efficiency region. The curves show that, for the pressure ratio range investi­gated, limiting loading was not reached.

Mass flow EW 6 / 6 is plotted against rotor exit-static- to inlet-total-pressure ratio P~/P; in figure 13 fo r lines of constant speed. At design pressure ratio (0.811) and speed, the mass flow was 1.477 pounds per second (0.670 kg/sec). This value is 4 percent lower than the design value of 1.537 pounds pe r second (0.697 kg/sec). This deficiency in mass flow is again attributed to the throat a r e a s of all s ta tors and ro tors being somewhat smaller than design. The variation of mass flow with pressure ratio for lines of constant speed is typical of turbines of subsonic design.

equivalent ­-design speed

0 120 - 0 loo -

0 90 D 70 -

- a 50 n 30

1.2

. 5'r 1.1

.4t. 4 '1% i , I . I .85 .90 , 5 Equiv ?ntexit-stat ic- t I pressure ratio, pglpi

F igure 13. -Var ia t i on of mass flow w i th pressure rat io and speed for two-stage operation.

19

I

0 20 40 60 80 100 120 Percent of equivalent design speed

Figure 14. -Va r ia t i on of torque wi th speed and pressure rat io for two-stage operation.

The variation of torque T E /6 with speed and p res su re ratio is shown in figure 14. A torque value of 104.44 inch-pounds (11.80 newton-meters) was obtained at design speed and at a pressure ratio of 1.233. This torque value is 3.5 percent lower than the design value of 108. 12 inch-pounds (12.22 newton-meters). Since specific work was equal to design value and the m a s s flow was 4 percent lower than design, the torque would be expected to be smaller than design. The slight upward trend of the torque-speed curves as speed is reduced is typical of curves obtained for two-stage axial-flow turbines. A zero-speed torque value of 23 1.1 inch-pounds (26.11 newton-meters) was obtained a t design pressure ratio. This torque is 2 .2 t imes that obtained at design speed and p res ­s u r e ratio.

Figure 15 shows the performance of the turbine from inlet to rotor exit in t e r m s of static and total efficiencies. Figure 15(a) indicates that design static efficiency

T s , 1to 5 of 0.825 was obtained at design speed and at a design blade-jet speed ratio of 0.465. The curves show that the efficiency is not affected by speed for speeds over 90 percent of design. Figure 15 (b) shows that a total efficiency vt, to of 0.845 was obtained at design speed and design blade-jet speed ratio. This value is only slightly lower than the design value of 0.850.

Figure 16 shows the overall turbine performance which includes the turbine exit collector. Figure 16 (a) indicates that design static efficiency to of 0.826 was obtained at design speed and design blade-jet speed ratio. Figure 16(b) indicates that the associated total efficiency qt,l to was 0.835 at design speed and design blade-jet

20

I

,,sign vi

I I / (a1 Static efficiency.

. s

. E

. I

. t

c .,

.4 Desian v

7 .;I

0 . I . 2

1-7

Percent of equivalent

design speed'i 0 120 4 0 100

Blade-jet speed ratio, u

(b) Total efficiency.

F igure 15. -Va r ia t i on of efficiency with blade-jet speed rat io for two-stage operation (based on rotor-exit conditions).

2 1

-

. 9

. 8

.D 0 - . 7 viF s:.6 equivalent -.-a, design speedV LL-a, 0 120 ­.-u . 5 0 100c m c m 0 90

D 70 . 4

Design value-, I 1 l ~ . .. 3 u I I

(a) Static efficiency.

t

Design value

I I 1 . 4 . 5

Blade-jet speed ratio, u

(bl Total efficiency.

Figure 16. - Variat ion of overal l eff iciency w i th blade-jet speed rat io for two-stage operation (based on collector-exit conditions).

speed ratio. This value of efficiency is 0.8 point lower than the design value of 0.843. Comparison of static efficiencies obtained at the rotor exit with those at the collector exit, for design speed and pressure ratio, shows that there was little static pressure recovery through the collector. In addition, comparison of total efficiencies shows that there was a drop of about 1point in total efficiency through the collector. These experi­mental values are the same as the design values, and hence the collector performed as designed.

Figure 17 shows the variation of tip static pressure through the turbine at design

22

1 2 3 4 5 Station

Figure 17. -Va r ia t i on of t ip static pressure th rough turb ine at equivalent design speed and design pressure ratio.

speed and pressure ratio. Comparison of the experimental static pressure variation with design shows that approximately design reaction was obtained ac ross all blade rows.

For two-stage operation at design speed and pressure ratio, the first-stage was operating at a rotor-exit-tip-static- to inlet-total-pressure ratio of 0.889. With this pressure ratio and the results from first-stage and two-stage performance data, the work split for two-stage operation was 3.15 Btu per pound (7.33 joules/gram) for the first stage and 2.81 Btu pe r pound (6.55 joules/gram) fo r the second stage, which repre­sents a stage work split of 53 to 47. The turbine was designed for an equal work split of 2.98 Btu pe r pound (6.937 joules/gram) pe r stage. Based on the two-stage total efficiency (based on rotor-exit conditions) of 0.845 and the measured first-stage total efficiency of 0.864, the associated second-stage total efficiency was computed to be 0.805, which is approximately 4 points lower than design.

The radial variation of rotor-exit total pressure and flow angle for two-stage operation is shown in figure 18 for design speed and design total- to static-pressure ratio. The trends shown in this figure are similar to those obtained at the first-stage rotor exit, as described in figure 11, with the general underturning at the exit again indicated.

23

c

x

I I I I I I I I I I Data

0 Experimental 0 Based o n wall static pressure

(a1 Exit total pressure

a--m m 30-c

cW

72 .76 .80 .84 .88 . 92 .96 I. 00 Radius ratio, r/router wall

(bl Exit flow angle.

f i g u r e 18. -Var ia t ion of rotor-exit total pressure and flow angle w i th radius rat io for two-stage operation at equivalent design speed and design pressure ratio.

SUMMARY OF RESULTS

An experimental investigation of a two-stage axial-f low turbine with a mean diameter of 8. 50 inches (21. 59 cm) was conducted to determine the performance a t design Reynolds number. This turbine was designed to drive a low-speed alternator for a 10-kilowatt­shaft-output space power system. Performance characterist ics were obtained for both first-stage and two-stage operation. The results of this investigation a r e summarized as follows:

1. An equivalent specific work of 5.96 Btu pe r pound (13.88 joules/gram) was obtained for two -stage operation at design speed and inlet -total - to rotor -exit -static -pressure ratio. The associated static and total efficiencies (based on turbine-inlet and rotor-exit conditions) were 0. 825 and 0. 845, respectively, which compare favorably with design.

2. A t this design point the flow was 1.477 pounds p e r second (0.670 kg/sec). This value was 4 percent lower than design and resulted principally from the flow areas in the blade rows being smaller than design since approximate design reaction was obtained.

3. Two-stage performance based on turbine-inlet and collector- exit conditions indi­cated static and total efficiencies of 0.826 and 0.835, respectively. These results indi­

24

cate a drop of approximately 1 point in total efficiency within the collector with no change in static efficiency and again agree closely with design.

4. Results of the first-stage tests indicated static and total efficiencies of 0.827 and 0. 867, which were approximately 2 points greater than design. This improved f i r s t -stage performance resulted in an indicated second-stage total efficiency of 0. 805, which is approximately 4 points less than design. As a result, a stage work split of 53 to 47 was indicated compared with a design equal split.

Lewis Research Center, National Aeronautics and Space Administration,

Cleveland, Ohio, September 18, 1967, 120-27-03 -13 -22.

25

I

APPENDIX - SYMBOLS

A

g

Ah

J

N

P

R

R e

r

T

U

V

W

W

CY

Y

6

E

71,

qt

'cr

flow area in. (cm2

gravitational constant, 32.174 ft/sec 2

specific work, Btu/lb (joules/gram)

mechanical equivalent of heat, 778.029 (ft-lb)/Btu

turbine speed, rpm

absolute pressure, psi (newton/cm 2)

gas constant (ft -lb)/(lb) (OR)(newton-meter/(kg) (OK))

Reynolds number, w/prm

radius, ft (m)

absolute temperature, OR (OK)

blade velocity, ft/sec (m/sec)

absolute gas velocity, ft/sec (m/sec)

ideal jet speed corresponding to total- to static-pressure ratio ac ross turbine, ft/sec (m/sec)

relative gas velocity, ft/sec (m/sec)

mass flow, lb/sec (kg/sec)

absolute gas flow angle measured from axial direction, deg

ratio of specific heats

ratio of inlet-total pressure to U. S. standard sea-level p re s su re

function of y used in relating parameters to those using air inlet conditions at U. S. standard sea-level conditions,

Y*-Y -1-1

static efficiency (based on inlet-total- to exit-static-pressure ratio)

total efficiency (based on inlet-total- to exit-total-pressure ratio)

squared ratio of critical velocity at turbine inlet to cri t ical velocity at U. S. standard sea-level temperature, (Vcr/V&) 2

26

I

1

2

3

4

5

6

p gas viscosity, lb/(ft)(sec) (Kg/(m)(sec))

I/ blade-jet speed ratio, U,/Vj

T torque, in. -1b (newton-meter)

Subscripts:

c r condition corresponding to Mach 1

eq air equivalent (U. S. standard sea level)

id ideal

m mean radius

station at turbine inlet

station at first-stage stator exit

station at first-stage rotor exit

station a t second-stage stator exit

station at second-stage rotor exit

station a t exhaust -pipe flange

Superscripts:

' absolute total state

* U. S. standard sea-level conditions (temperature, 518.67' R (288. 15' K); pressure, 14.696 psia (10.128 newton/cm 2))

27

REFERENCES

1. Bernatowicz, Daniel T. : NASA Solar Brayton Cycle Studies. Paper presented at the Symposium on Solar Dynamics Systems, Interagency Advanced Power Group, Washington, D. C. , Sept. 24-25, 1963.

2. Kofskey, Milton G; and Holeski, Donald E. : Cold Performance Evaluation of a 6.02-Inch Radial Inflow Turbine Designed for a 10-kilowatt Shaft Output Brayton Cycle Space Power Generation System. NASA TN D-2987 (Rev. ) 1966.

3. Holeski,, Donald E. ; and Futral, Samuel M. , Jr.: Experimental Performance Evalua­tion of a 6.02-Inch Radial-Inflow Turbine Over a Range of Reynolds Number. NASA TN D-3824, 1967.

4. Cohen, R. ; Gilroy, W. K. ; and Havens, F. D. : Turbine Research Package for Research and Development of High Performance Turboalternator. Rep. No. PWA -2796 (NASA CR-54885), Pratt and Whitney Aircraft, Jan. 1967.

5. Forrette, Robert E. ; Holeski, Donald E. ; and Plohr, Henry W. : Investigation of the Effects of Low Reynolds Number Operation on the Performance of a Single-Stage Turbine with a Downstream Stator. NASA TM X-9, 1959.

6. Cohen, R. ;Gilroy, W. K. ; and Havens, F. D. : Turbine Research Package for Research and Development of High Performance Axial Flow Turbine-Compressor. Rep. No. PWA-2822 (NASA CR-54883), Prat t and Whitney Aircraft, Dec. 1966.

28 NASA-Langley, 1968 -3 E-3 983

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